1. Field of the Invention
Embodiments of the present invention relate generally to compressors and, more specifically, to a mid-span gas bearing in a multistage compressor.
2. Description of the Prior Art
A compressor is a machine which increases the pressure of a compressible fluid, e.g., a gas, through the use of mechanical energy. Compressors are used in a number of different applications and in a large number of industrial processes, including power generation, natural gas liquification and other processes. Among the various types of compressors used in such processes and process plants are the so-called centrifugal compressors, in which the mechanical energy operates on gas input to the compressor by way of centrifugal acceleration, for example, by rotating a centrifugal impeller.
Centrifugal compressors can be fitted with a single impeller, i.e., a single stage configuration, or with a plurality of centrifugal stages in series, in which case they are frequently referred to as multistage compressors. Each of the stages of a centrifugal compressor typically includes an inlet volute for gas to be compressed, a rotor which is capable of providing kinetic energy to the input gas and a diffuser which converts the kinetic energy of the gas leaving the impeller into pressure energy.
A multistage compressor 100 is illustrated in FIG. 1. Compressor 100 includes a shaft 120 and a plurality of impellers 130-136 (only three of the seven impellers are labeled). The shaft 120 and impellers 130-136 are included in a rotor assembly that is supported through bearings 150 and 155.
Each of the impellers 130-136, which are arranged in sequence, increase the pressure of the process gas. That is, impeller 130 may increase the pressure from that of gas in inlet duct 160, impeller 131 may increase the pressure of the gas from impeller 130, impeller 132 may increase the pressure of the gas from impeller 131, etc. Each of these impellers 130-136 may be considered to be one stage of the multistage compressor 100.
The multistage centrifugal compressor 100 operates to take an input process gas from inlet duct 160 at an input pressure (Pin), to increase the process gas pressure through operation of the rotor assembly, and to subsequently expel the process gas through outlet duct 170 at an output pressure (Pout1) which is higher than its input pressure. The process gas may, for example, be any one of carbon dioxide, hydrogen sulfide, butane, methane, ethane, propane, liquefied natural gas, or a combination thereof.
The pressurized working fluid within the machine (between impellers 130 and 136) is sealed from the bearings 150 and 155 using seals 180 and 185. A dry gas seal may be one example of a seal that can be used. Seals 180 and 185 prevent the process gas from flowing through the assembly to bearings 150 and 155 and leaking out into the atmosphere. A casing 110 of the compressor is configured so as to cover both the bearings and the seals, and to prevent the escape of gas from the compressor 100.
While additional stages can provide an increase in the ratio of output pressure to input pressure (i.e. between inlet 160 and outlet 170), the number of stages cannot simply be increased to obtain a higher ratio.
An increase in the number of stages in a centrifugal compressor leads to multiple problems. The bearings which support the shaft are outside a sealed area that includes the impellers. An increase in the number of stages necessitates a longer shaft. A longer shaft cannot be safely supported by the bearings for the same operating speed, which become further apart as the shaft length increases making the shaft more flexible.
As the rotor assembly gets longer, the shaft becomes flexible therefore decreasing the rotor natural frequencies. When operating at higher speeds, the decrease in the fundamental natural frequencies of the rotor assembly tends to make the system more susceptible to rotor-dynamic instability, which can limit the operating speed and output of the machine.
The other issue is the forced response due to synchronous rotor imbalance. When the operating speed coincides with a rotor natural frequency, the machine is defined to be operating at a critical speed, which is a result of rotor imbalance. The compressor must pass through several of these natural frequencies or critical speeds before reaching the design operating speed.
As the compressor passes through critical speeds, the vibration amplitude of the rotor must be bounded by damping from bearings. However, with a long shaft, the majority of the rotor-dynamic energy is transferred to bend the rotor instead of energy dissipation at the bearings. This results in low damped rotor modes and high amplification factors at rotor resonances that can lead to casing and impeller rubs and even catastrophic failure of the machine.
At higher speeds past the rotor critical speeds, fluid induced forces are generated between the rotor assembly and the casing (i.e. fluid induced rotor dynamic instability). These pulsations, stemming from fluid forces can excite destructive or even catastrophic vibrations if not adequately dampened. Rotor-dynamic instability is a different mechanism from critical speeds or imbalance response and often time is much more difficult to address.
It would be desirable to design and provide a multistage centrifugal compressor which includes additional stages without increasing the diameter of the shaft and other design parameters that would drastically change the size and cost of the machine.